Motor having single cone fluid dynamic bearing balanced with shaft end magnetic attraction

ABSTRACT

A single cone fluid dynamic bearing motor, including a shaft having a diminishing conical taper surface, a sleeve having a concavity opposite the shaft, lubricant filled in a clearance between the shaft and the sleeve, and magnetic members to generate magnetic attraction between one end of the shaft and a cone apex of the sleeve. Grooves are formed on the conical taper surface of the shaft or the sleeve so as to create load capacity when the motor rotates, whereby rotating parts of the motor are supported by the axial components of the load capacity balanced with the magnetic attraction. The motor thereby achieves reduction in thickness, current, and cost, and inhibits non-repeatable runout.

BACKGROUND OF THE INVENTION

[0001] 1. Technical Field of the Invention

[0002] The present invention relates to a fluid dynamic bearing motor,and more particularly to a fluid dynamic bearing having a conical shapeto enable the motor to be smaller in thickness and lower in cost.

[0003] 2. Description of the Related Art

[0004] There has been a trend toward the fluid dynamic bearing motor asthe power source for rotary memory devices, cooling fans, and the like,because of its quietness in operation and the necessity to reducenonrepeatable runout (NRRO) of rotating parts. Portable applications ofsuch electronic devices have been widespread, increasing the demands forfurther reduction in their thickness and required current. However,there are limitations on further reduction in thickness of the fluiddynamic bearings, because they need to have a certain span between thebearings for supporting the shaft in order to inhibit NRRO. Also, inorder to maintain a constant clearance between the bearings, they mustbe machined with extreme precision in the order of submicrons, wherebyit is difficult to produce them at low cost.

[0005] In order to make fluid dynamic bearings thinner, a novelstructure is necessary which does not require two bearings forsupporting the shaft at axially spaced positions. The bearings shouldhave as little sliding area as possible so as to achieve a reduction inthe required current. Further, cost reduction will be achieved throughthe development of a structure wherein the bearing clearance ismaintained with necessary accuracy even with the components machinedwith a lower degree of precision.

[0006] Single cone fluid dynamic bearings, which can support loads ofboth radial and thrust directions, have attracted attention as havingpotentialities in many respects. However, while some single conestructures that help decrease the thickness of the bearing have beenproposed, for example, in Japanese Laid-open Utility Model PublicationNo. Hei. 06-004731, these are for air dynamic bearings and anyway havenot been very successful. The main reason is that the single conebearing is structurally incapable of sufficiently inhibiting NRRO duringrotation. Japanese Laid-open Patent Publications No. 2000-004557 and No.2000-205248 propose combined use of a conical bearing and a cylindricalbearing to improve the overall performance. However, the cylindricalbearing requires high-degree machining precision for maintaining aconstant bearing clearance, canceling out the advantages of the conicalbearing. U.S. Pat. No. 5,854,524 discloses a single semi-spherical airdynamic bearing having a similar structure as that of the single conebearing, but in this case also, the radius of two spherical surfacesmust be strictly controlled to secure a sufficient radial load capacity,because of which cost reduction is hardly achievable.

[0007] Thus the problems yet to be resolved in single cone fluid dynamicbearing motors are how to improve the stability in its rotatingattitude, and how to realize a structure which prevents leakage of thelubricant and yet is easy to assemble.

SUMMARY OF THE INVENTION

[0008] An object of the present invention is to resolve these problemsand to provide a single cone fluid dynamic bearing motor which can bereduced in thickness and required current, and is simple and can beproduced at lower cost.

[0009] A fluid dynamic bearing motor according to the invention includesa shaft having a diminishing conical taper surface, a sleeve having aconical concavity opposite the shaft, lubricant filled in a clearancebetween the shaft and the sleeve, and means for generating magneticattraction between one end of the shaft and a cone apex of the sleeve,and an annular wall arranged around the shaft to face an outercircumferential wall of the sleeve, a clearance between the annular walland the outer circumferential wall of the sleeve being increased inwidth toward an open end to form a taper seal of the lubricant. In thisconstruction, a plurality of grooves are formed on a conical tapersurface of one of the shaft and the sleeve, and the grooves are providedfor creating load capacity when the motor rotates, whereby rotatingparts of the motor are supported by axial components of the loadcapacity balanced with the magnetic attraction. The boundary of thelubricant is positioned around the sleeve, so as to enable a reliableseal to be formed even in high-speed operation.

[0010] A ring-shaped member is fixed to one end of the annular wallwhich is arranged around the shaft, and an annular recess is provided inthe outer circumferential wall of the sleeve, the inner periphery of thering-shaped member being positioned within the annular recess, so as torestrict an axial movable distance of the rotating parts. This structureserves as a stopper for the rotating parts in the case where the motoris subjected to a large shock.

[0011] The means for generating magnetic attraction includes a permanentmagnet and a magnetic material, respectively provided inside the shaftand in the apex of the sleeve opposite to the shaft, or vice versa.Magnetic attraction developed at the end of the shaft acts on the shaftto adjust its position in cooperation with the load capacity created bythe grooves, thereby ensuring the stable attitude of rotating parts.

[0012] Moreover, the following structures for a permanent magnet toprotrude from one end of the shaft are proposed. The shaft includes apermanent magnet held inside. The permanent magnet is assembled with theshaft such that it is initially held movably but firmly enough toovercome the magnetic attraction as being substantially protruded fromone end of the shaft, and is pressed into the shaft by a pressure largerthan the magnetic attraction applied from both ends of the shaft and thesleeve to a predetermined position, where the cone apex of the sleeve ora plate spring interposed between the apex of the sleeve and thepermanent magnet is resiliently deformed, whereby when the motor isstationary the permanent magnet and the apex of the sleeve or the platespring make contact with each other, while they are brought out ofcontact when the motor is rotating, by a distance equal to or shorterthan an axial flying height determined on conical surfaces of the shaftand sleeve. Thereby, the start-up failure caused by the conical surfaceof the shaft being fitted in the sleeve when the motor is not inoperation can be avoided, improving the reliability of the motor.

[0013] Alternatively, the grooves may be formed on both opposite tapersurfaces of the shaft and the sleeve at the almost same axial positions.In this constitution, the grooves have different angular length fromeach other in the circumferential direction. Thereby, each delay, fromthe time point when the bearing clearance becomes small until the timepoint when the pressure in the lubricant in the clearance becomes localmaximum by the corresponding groove, is varied in proportion to thecorresponding angular length of each of the grooves. Thereby, animproved constitution which can avoid half whirls and other unstablemovements of the motor can be provided.

[0014] Alternatively, conductive magnetic powder may be mixed in thelubricant so as to electrically connect the rotating parts to groundbecause of the powder gathered in the magnetic field between the end ofthe shaft and the apex of the sleeve.

[0015] According to the fluid dynamic bearing motor of the presentinvention, the load capacity created by the rotation of the motor actsvertically with respect to the conical surfaces, causing the shaft andthe sleeve to rotate in non-contact relationship at a position where theaxial components of the load capacity and the magnetic attraction are inequilibrium. The radial components of the load capacity counterbalanceeach other at respective circumferential points, thereby contributing tothe centering of the rotating parts. The load capacity itself actsvertically on the tapered surface of the cone, and therefore it servesto adjust the attitude of rotating parts when they tilt with respect tothe fulcrum conforming to the cone apex.

[0016] Magnetic attraction developed at the end of the shaft acts on theshaft to adjust its position in cooperation with the load capacitycreated by the grooves, thereby ensuring the stable attitude of rotatingparts.

[0017] The main reason why the prior art single cone bearing has failedto maintain the attitude of rotating parts is that the bearing wasprovided only with a load equal to the weight of its own, or even lessthan that by using a magnetic bearing in order to avoid friction duringthe initial and final periods of operation as disclosed in JapaneseLaid-open Utility Model Publication No. Hei. 06-004731. As has beenexplained above, a good balance is achieved between two forces of theaxial component of load capacity of the bearing versus the load.Therefore, a small load can only create a small load capacity, which isinsufficient to create forces for maintaining stable attitude ofrotating parts. In the fluid dynamic bearing of the present invention, alarge load is applied on the bearing by the magnetic attraction actedbetween the shaft and the sleeve. Therefore, the load capacity of thebearing, which counterbalances the load, can be set to a desired largevalue, whereby the stability of the attitude of rotating parts isimproved. The magnetic attraction may be varied case by case dependingon permissible level of NRRO, the size of the motor, and various otherconditions.

BRIEF DESCRIPTION OF THE DRAWINGS

[0018] These and other objects and advantages of the present inventionwill become clear from the following description with reference to theaccompanying drawings, wherein:

[0019]FIG. 1 is a cross sectional view showing a fluid dynamic bearingmotor according to an embodiment of the present invention;

[0020]FIG. 2 illustrates the means for generating magnetic attractionand the lines of magnetic flux in the embodiment shown in FIG. 1;

[0021]FIG. 3(a) and FIG. 3(b) illustrate the bearing section in detail,FIG. 3(b) being a plan view of a sleeve, and FIG. 3(a) being a crosssectional view of a shaft and the sleeve;

[0022]FIG. 4(a) illustrates a cross-section of the shaft and the sleeve,and component forces of load capacity, and FIG. 4(b) illustrates thedistribution of pressure developed during the rotation;

[0023]FIG. 5(a) shows an embodiment in which magnetic attractiongenerating means is provided at one end of the shaft, and FIG. 5(b)shows another embodiment using a rotor magnet as the magnetic attractiongenerating means, given in explanation of the difference in how aposition adjusting force acts on the rotary section;

[0024]FIG. 6 is a detailed cross sectional view of the bearing sectionhaving a permanent magnet at one end of the shaft for limiting contactbetween the shaft and the sleeve when they are stationary;

[0025]FIG. 7 is an explanatory view illustrating how the permanentmagnet of FIG. 6 is fitted in a predetermined position;

[0026]FIG. 8(a) illustrates a cross-section of the bearing sectionhaving a crown, with a graph showing the pressure distribution, and FIG.8(b) illustrates how the load capacity acts on the rotary section whenit is offset from the center;

[0027]FIG. 9(a) illustrates the pressure distribution with across-section of the bearing section having a crown and spiral grooves,and FIG. 9(b) illustrates how the load capacity acts on the rotarysection when it is offset from the center;

[0028]FIG. 10(a) and FIG. 10(b) are detailed views of the bearingsection having a modified construction wherein grooves are formed onboth opposite surfaces of the shaft and the sleeve, FIG. 10(a) being aplan view of the sleeve, and FIG. 10(b) being a cross sectional view ofthe shaft and the sleeve.

[0029]FIG. 11 is a cross sectional view of a modified construction ofthe embodiment in which a channel is formed through the shaft;

[0030]FIG. 12(a) and FIG. 12(b) are explanatory views illustrating how aring-shaped member can be axially adjusted, FIG. 12(a) being an enlargedcross sectional view of the vicinity of the ring-shaped member, and FIG.12(b) being a cross sectional view of the bearing section;

[0031]FIG. 13 illustrates major parts of the bearing section having amodified construction in which is used conductive magnetic powder; and

[0032]FIG. 14 is a cross sectional view of the prior art fluid dynamicbearing motor.

DESCRIPTION OF THE PREFERRED EMBODIMENT

[0033] Preferred embodiments of a fluid dynamic bearing motor accordingto the present invention will be hereinafter described with reference tothe accompanying drawings.

[0034] The prior art fluid dynamic bearing motor structure is reviewedby FIG. 14 before the description of present invention. The fluiddynamic bearing motor possesses two radial bearings which are providedon the surface of shaft 91 or cylindrical sleeve 92, and two thrustbearings which are provided on the surface of both sides of a thrustplate 93, and has herringbone grooves respectively in each bearing. Theclearance between thrust plate 93, sleeve 92, and thrust bush 94 whichcompose thrust bearings are ten-micron meter level, and also theclearance between shaft 91 and sleeve 92 which compose radial bearingsis two-micron meter level with the lubricant.

[0035] Two radial bearings and the existence of thrust plate 93 make theentire motor thin difficult. The bearing clearance and also the rightangle degree of shaft 91 and hub 95, shaft 91 and thrust plate 93 shouldbe well controlled at the mass production stage because the loadcapacity of the bearing depends on the clearance. These are factors ofincreasing the cost. Moreover the joint part of the thrust bush 94 andthe sleeve 92 in a portion which the lubricant contacts, is joined toprovide a seal by swaging, bonding, or laser welding. The lubricantleakage may be caused from the joint part space and a serious trouble isoften invited. Reference numerals 96, 97, 98, and 99 respectivelyrepresent a rotor magnet, a stator core, coils, and a base.

[0036]FIG. 1 is a cross sectional view of a fluid dynamic bearing motoraccording to a embodiment of the present invention. A shaft 11 has adiminishing conical taper, and a sleeve 12 arranged opposite the shaft11 has a conical concavity. The clearance between the shaft 11 and thesleeve 12 is filled with magnetic oil as the lubricant. The shaft 11 issurrounded by an annular wall 23, and the clearance between the annularwall 23 and the outer circumference of the sleeve 12 becomes wider in anaxial direction, thereby forming a taper seal, where there is theboundary 17 of the lubricant. A permanent magnet 35 is provided within athrough hole 39 in the shaft 11, so as to generate magnetic attractionbetween itself and the top of the sleeve 12 which is made of a magneticmaterial.

[0037] The permanent magnet 35 is fitted within the through hole 39 suchthat it is initially held with clearance so as to be movable as beinglargely protruded from the shaft 11. It is then brought into contactwith the inside top limit of the sleeve 12, and is adjusted via thethrough hole 39 and fixed in position at the time of assembling suchthat the permanent magnet makes contact with the sleeve when the motoris stationary, while they are brought out of contact when the motor isrotating, by a distance equal to or shorter than an axial flying heightdetermined on conical surfaces of the shaft and sleeve.

[0038] Rotary section is composed of the shaft 11, the annular wall 23,a hub 41, a rotor magnet 44, and others, and fixed section is composedof the sleeve 12, a base 43, a stator core 47, coils 50, and others.

[0039] The bearing section is constituted by the shaft 11, the sleeve12, and a series of herringbone grooves, to be described later, providedin one of the conical taper surfaces 13 of the shaft 11 and the sleeve12. The grooves serve to pump the lubricant toward their center toincrease the pressure of the lubricant. The load capacity therebycreated is in inverse proportion to the size of the clearance betweenthe shaft and sleeve. Therefore, the clearance size is determined suchthat the axial components of the load capacity and the above-mentionedmagnetic attraction are in equilibrium, while radial components of theload capacity are used for the centering of the shaft 11. Accordingly,the magnetic attraction, which determines the load capacity, is set sothat the load capacity is large enough to support the rotary sectionduring rotation. The clearance, accordingly, is approximately severalmicrometers wide. When the apical conical angle of the bearing sectionis large, the axial components of the load capacity may be given moreconsideration, while the radial components play a more important rolewhen the apical conical angle is small. In this embodiment, the angle ofthe cone apex is slightly smaller than 60° so as to give more weight tothe radial components to ensure precise centering of the shaft.

[0040] The stator core 47 and the coils 50 cooperate with the rotormagnet 44 to drive the rotary section. The rotary section furtherincludes a magnetic or optical disk or the like carried thereon as aload. The force applied to the interface between the shaft 11 and thesleeve 12 varies depending on the manner in which the memory device isinstalled in a normal state or inverted state. That is, if the device isset in a normal state, the bearing receives the weight of the movableparts in addition to the magnetic attraction. If the device is set in aninverted state, the bearing receives a load less than the magneticattraction because the weight of the movable parts is subtractedtherefrom. In light of this, the magnetic attraction should beapproximately three times larger than the weight of the movable parts,which has empirically been confirmed to ensure stable rotating attitudeof the rotary section. If the magnetic attraction is increased so as tocreate accordingly larger load capacity, precession of the shaft canfurther be restricted and its attitude can be made more stable. On theother hand, it has been ascertained that such increase in the magneticattraction causes the sliding friction to become larger at the time ofstarting up or stopping the motor, resulting in shorter operable life ofthe bearing. Therefore, in the case of the fluid dynamic bearing motorfor a small magnetic disk device, magnetic attraction should beapproximately five times larger than the weight of the movable parts,which is the sum of the weight of the rotary section and the loadweight. Such settings may be determined case by case depending on therequired precision for the rotating attitude of rotary section.

[0041]FIG. 2 illustrates a cross-section of the shaft 11, permanentmagnet 35, sleeve 12 and others, together with lines of magnetic flux.The shaft 11 is made of a non-magnetic material while having thepermanent magnet 35 inside, which is made of highly magnetic rare earth.Provided that the bearing section has about 5 mm diameter, the designvalue allotted to the diameter of the permanent magnet 35 is 1 to 2 mm.Since the top of the shaft 11 and the sleeve 12 are arranged with a verysmall clearance of about 10 micrometers therebetween, magneticattraction remains constant irrespective of the variations in thisclearance. Thus the tolerance in machining and assembling can be setlarger. Reference numeral 55 indicates the direction of magnetization ofthe permanent magnet 35. The distal end of the permanent magnet 35 isformed spherical so as to concentrate the magnetic flux. The magneticflux 56 thus intensified enters the truncated cone top end of the sleeve12, and as shown by the reference numerals 57 and 58, it passes throughthe sleeve 12 and returns to the other end of the permanent magnet 35via the conical tapered surface. The magnetic flux 58 from the conicaltapered surface to the end of the permanent magnet 35 flies a longdistance and over a large area and thus is distributed and low inintensity. Accordingly, the magnetic attraction between the conicaltapered surface of the sleeve 12 and the permanent magnet 35 is weak.

[0042] In this specific example, magnetic oil is used as the lubricant.Therefore, a centripetal force acts on the magnetic oil around thetruncated conical top of the sleeve 12 where the magnetic intensity ishigh, whereby air bubbles are positively eliminated from the bearingsection. Normal oil can also be used for achieving the intended purposeof the bearing section.

[0043] Apart from the structure shown in FIG. 1, magnetic attractioncould be developed using the rotor magnet 44 and the stator 47 axiallyoffset from each other, or the rotor magnet 44 and the magnetic piecearranged below the rotor magnet. However, the former has a disadvantagethat it produces vibration, and the latter causes an increase inconsumed current because of the Foucault current developed in themagnetic piece. The magnetic attraction generating means in thisembodiment can resolve all these problems encountered by theabove-mentioned other mechanisms.

[0044]FIG. 3(a) and FIG. 3(b) illustrate the structure of the bearingsection of the embodiment shown in FIG. 1 in more detail. FIG. 3(b) is aplan view of the sleeve 12, and FIG. 3(a) is a cross sectional view ofthe shaft 11 and the sleeve 12. As shown in FIG. 3(b), a series ofherringbone grooves 18 is provided on the taper surface 13 of the sleeve12. The grooves 18 are V-shaped shallow recesses of about severalmicrometers depth. When the motor rotates, the grooves 18 pump thelubricant from the outer and inner peripheral sides toward their centralpointed ends to increase the pressure of the lubricant, so as to liftthe shaft 11 from the sleeve 12 and support it in a flying state. Inthis embodiment, the grooves are formed so that the pumping capacityfrom the outer peripheral side toward the inner peripheral side islarger than that from the inner peripheral side toward the outerperipheral side, whereby the pumping capacity towards the innerperipheral side remains and the pressure of the lubricant on the innerperipheral side can be increased swiftly when starting up the motor, soas to decrease the sliding friction between the shaft 11 and the sleeve12. The grooves 18 illustrated in FIG. 3(b) have larger groove length onthe inner peripheral side, but this does not contradict the descriptionin the foregoing, since the pumping capacity is determined by thediminishing degree of the circumferential length of the grooves and theradial length of the grooves.

[0045] The clearance between the annular wall 23 and the outercircumference of the sleeve 12 becomes wider in an axial direction,where a taper seal is formed, which provides a seal by the surfacetension of the lubricant. To one end of the annular wall 23 is fixed aring-shaped member 24, of which inner periphery fits in an annularrecess 26 formed on the outer circumferential wall of the sleeve 12,thereby restricting displacement of the rotary section in axialdirections. The ring-shaped member 24 is either resilient or partiallycut out so as to be rotatably fitted into the annular recess 26 inadvance during the assembly of the bearing components. Thereafter, thering-shaped member 24 is fixed to the end face of the annular wall 23 byspot-welding or bonding through access holes 25. Three such access holes25 are provided at circumferentially spaced points so as to evenlysecure the ring-shaped member 24.

[0046] Since the taper seal of the lubricant is formed not on the outerperiphery of the conical bearing surface but on the outer circumferenceof the sleeve 12, the overall thickness of the motor can be madesmaller. Meanwhile, the taper seal can have a sufficient space in theaxial direction, whereby the taper angle can be made as small as 10° orlower to form a strong seal of the lubricant. The boundary 17 of thelubricant is therefore formed not between conical surfaces, but betweensubstantially vertical outer walls of the sleeve 12 and the annular wall23. Therefore there is no risk that the lubricant may leak undercentrifugal force even in high-speed operation.

[0047]FIG. 4(a) and FIG. 4(b) illustrate the distribution of pressuredeveloped in the lubricant when the motor rotates and the componentforces of the load capacity applied to the interface between the shaft11 and the sleeve 12 in accordance with the pressure distribution. Thesedrawings are given in explanation of how the rotating attitude of theshaft is self-adjusted.

[0048]FIG. 4(b) shows various features 62, 63, 64, and 65 of thepressure distribution of the lubricant caused by the grooves 18 inoperation. The y-axis 60 represents pressure, while the x-axis 61indicates radial coordinates corresponding to FIG. 4(a). The pressurereaches a highest point 63, 65 at positions substantially correspondingto the pointed ends of the V-shaped grooves 18. The drawing shows thepressure distribution without the influence of the atmospheric pressure,and therefore the pressure 62 at an outer peripheral point is almostzero. On the other hand, the pressure 64 at an inner peripheral point ishigher than the atmospheric pressure, because the grooves 18 are formedto have larger pumping capacity towards the inner peripheral side.

[0049]FIG. 4(a) shows a cross-section of the shaft 11 and the sleeve 12.Reference numerals 67, 68 in FIG. 4(a) represent the load capacitycreated as the pressure in the lubricant increases. It should be notedthat such a load capacity is created at each one of the severalcircumferentially located points, but only two of these are shown in across-section for the ease of explanation.

[0050] Reference numerals 69, 71 represent the axial components of theload capacity 67, 68, respectively. Reference numerals 70, 72 representrespective radial components thereof. Since the load capacity 67, 68 issubstantially in inverse proportion to the size of the clearance betweenthe shaft 11 and the sleeve 12, the clearance is determined such thatthe axial components 69, 71 and the magnetic attraction between therotary section and the fixed section are in equilibrium. The radialcomponents 70, 72 act in opposite directions so that they counterbalanceeach other, whereby the shaft 11 is centered.

[0051] The load capacity 67, 68 acts vertically to the conical surfaces.Thus, it acts on the shaft 11 as moment, i.e., the distance L multipliedby the load capacity 67, 68, where L is the distance from an imaginaryfulcrum 66 corresponding to the cone apex and the point from which theload capacity 67, 68 acts. The moment resulting from the load capacity67, 68 acts in reverse directions, and because the load capacity 67, 68is substantially in inverse proportion to the nearby clearance betweenthe shaft 11 and the sleeve 12, the moment caused by the load capacity67, 68 acts around the fulcrum 66 as a position adjusting force for theshaft 11, counterbalancing each other to equalize the clearance betweenthe shaft 11 and the sleeve 12. Thereby, the attitude of the shaft 11 ismaintained upright, and its precession is restricted.

[0052] Viscosity of the oil used as the lubricant generally decreases ata high temperature, leading to a decrease in the load capacity. It isthe practice in the prior art to set the load capacity high to allow forthe decrease in pressure over a maximum limit of the temperature rangefor use, as a result of which there are the problems of excessive loadcapacity and large current at lower temperatures. According to theinvention, the clearance between the shaft 11 and the sleeve 12 ischanged corresponding to the equilibrium between the axial components69, 71 of the load capacity 67, 68 and the magnetic attraction, andtherefore the load capacity is kept substantially constant irrespectiveof the temperature. That is, a temperature compensation is automaticallyprovided. This allows the load capacity to be set constant over theentire range of temperatures, eliminating the problems of excessive loadcapacity or current at low temperatures, and enabling a design with lowcurrent to be made.

[0053] Furthermore, the motor according to the invention is low inrespect of bearing loss. Bearing loss of the fluid dynamic bearing ismainly caused by friction between the surfaces of the shaft 11 andsleeve 12 and the lubricant in small clearances where the grooves exist.The bearing according to the invention has only a series of grooves,which is a practical minimum, and thereby can achieve a reduction inrequired current.

[0054] The moment which acts on the shaft 11 to maintain its attitude isdefined by the product which is obtained by multiplying the distance Lby the load capacity 67, 68 as noted above. Therefore, there is no needto provide two series of grooves with a large span therebetween in anaxial direction as in the prior art. The motor according to theinvention needs only one series of grooves 18, therefore the structureis more simple and thinner than the prior art.

[0055]FIG. 5(a) and FIG. 5(b) are given in explanation of how themagnetic attraction at one end of the shaft effectively acts to restoreits rotating attitude. FIG. 5(a) illustrates one example of theembodiment in which magnetic attraction is developed at one end of theshaft, and FIG. 5(b) illustrates another example having a rotor magnetas the magnetic attraction generating means. Both of these drawings showa state wherein the upper part of the shaft 11 is tilted leftwards, andthe load capacity created by the dynamic pressure is denoted atreference numerals 67, 68 on the left and right sides similarly to FIG.4(a). The load capacity 67 is larger than the load capacity 68 becauseof the difference in the bearing clearance, thereby acting as a momentforce on the shaft 11 to restore its rotating attitude as has beendescribed with reference to FIG. 4(a). It will be understood from FIG.5(a) that the magnetic attraction 83 acting from the top end of theshaft 11 serves as the moment to adjust the rotating attitude jointlywith the load capacity 67, 68. In the example shown in FIG. 5(b) usingthe rotor magnet 44 as the magnetic attraction generating means, themagnetic forces 84, 85 on both sides are substantially balanced witheach other, but these magnetic forces 84, 85 developed between the rotormagnet 44 and the magnetic piece 53 are in inverse proportion to theclearance therebetween. Therefore, the magnetic force 84 on the side onwhich the clearance is smaller becomes larger than the magneticattraction 85, disturbing the balance between the load capacity 67, 68.Thus the structure in which one end of the shaft has magnetic attractioncan more advantageously help maintain the stable attitude of rotarysection.

[0056]FIG. 6 and FIG. 7 are detailed views of the bearing sectionillustrating a modified construction of the embodiment wherein thepermanent magnet prevents the shaft and the sleeve from making surfacecontact with each other when they are stationary. As shown in FIG. 6,the permanent magnet 35 is provided at the top end of the shaft 11, suchas to contact a plate spring 33 placed at the inside top limit of theconical sleeve 12 when stationary. The dotted lines 11 a illustrate theposition of the shaft when stationary, while the solid lines indicatethe position of the shaft 11 when rotating. The permanent magnet 35protrudes by a predetermined amount such that f≧d, where d is thedistance between the top of the permanent magnet 35 and the inside toplimit of the sleeve 12, and f is the axial flying height of the shaft 11from the sleeve 12 measured at conical surfaces. To be specific, thepermanent magnet 35 is protruded so that f—d is about 5 micrometers ifthe flying height is within the range of 10 to 20 micrometers, takinginto account that the flying height f of the shaft 11 varies dependingon temperatures. Thus the top of the shaft 11 flies up from the platespring 33 at least about 5 micrometers during rotation, while itsconical surface flies up to an axial height of about 10 to 20micrometers, maintaining a stable rotating attitude.

[0057] Conical bearings have a potential risk that the shaft fits intothe sleeve, increasing the friction therebetween, resulting in start-upfailure. This is caused by various factors such as the intensity ofmagnetic attraction, the apical conical angles, and the hardness of thematerial making up the shaft and sleeve, correlating with each other.Small motors to which the present invention is applied are relativelyfree of such troubles, but the structure shown in FIG. 6 further ensuresthat no such troubles will occur.

[0058]FIG. 7 is given in explanation of how the permanent magnet shownin FIG. 6 is adjusted in position. The permanent magnet 35 is initiallyfitted in the cylinder 32 inside the shaft 11 with clearance so as to bemovable, but firmly enough to overcome the magnetic attraction. Forassembling the permanent magnet 35, it is placed upon the shaft 11 asbeing protruded substantially therefrom, and the sleeve 12 is coupledthereon. Pressure that is larger than the magnetic attraction is thenapplied to the sleeve 12 and the shaft 11 so that the permanent magnet35 contacts the plate spring 33 placed at the inside top limit of thesleeve 12, until the shaft 11 and the sleeve 12 make surface contactwith each other on their conical surfaces and the plate spring 33 isresiliently deformed. The dotted lines 11 b show the position of shaftunder the pressure and the dotted lines of plate spring 33 indicate thedeformed one under pressure, while the solid lines indicate the platespring 33 having restored initial shape, after the pressure has beenremoved. As the plate spring 33 resiliently returns into its initialshape, a clearance is created between the conical surfaces of the shaft11 and the sleeve 12. The resilient deformation of the top of the sleevemay be arranged on the inside top limit of the sleeve 12 instead ofutilizing a plate spring.

[0059] After the position alignment, the permanent magnet 35 shouldpreferably be fixed in position by bonding or welding, so as towithstand large shocks. Further, it is preferable to provideantifriction measures on the top of the permanent magnet 35 and theopposite plate spring 33 placed at inside top limit of the sleeve 12such as application of a ceramic material or plating treatment, so as toensure stable performance over a long time.

[0060] Single cone bearings have the characteristics that even when theshaft and the sleeve have slightly different diameters, they still canface each other at given axial positions, whereby the tolerance of theirdimensions can be made large, offering the advantage of lower cost. Thepermanent magnet 35 shown in FIG. 6 could initially be fixed to theshaft 11, but in that case the diameters of the shaft 11 and the sleeve12 and the protruding amount of the permanent magnet 35 must preciselybe controlled. If the demands for the performance of the fluid dynamicbearing motor in regard to inhibition of NRRO are relatively low, thensuch control of dimensions could easily be achieved, while it is not ifthe demands are high. Thus the total cost would be lower with thestructure wherein the permanent magnet allows itself to be positionallyadjusted as in this embodiment.

[0061]FIG. 8(a) shows a shaft fixed type embodiment in which hasherringbone grooves and crown in the conical surface. The herringbonegrooves in the conical surface are formed to have flat region in centralparts. While the grooves 20, 21 on both sides are shown in the crosssectional view so that their positions are more clearly understood, theyare actually formed on the surface of the conical shaft 11, having aseveral micrometers depth. The shaft 11 has a slightly bulging crown 19on its conical surface so as to have a flat band region where thebearing clearance is minimum. Correspondingly, a circumferential groove40 of about 10 micrometers depth is provided in the sleeve 12 oppositethe flat band region formed by the crown 19. Specific dimensions of thecrown 19 may differ case by case depending on various conditions, butbasically they are set such that the bearing clearance at the outermostperiphery of the conical shaft 11 and the sleeve 12 is severalmicrometers larger than that in the flat band region. With thisconstruction, even if the apical conical angles of the shaft 11 and thesleeve 12 are not precisely in conformity with each other, edge contactsat the inner and outer peripheries can be prevented. Therefore, themachining tolerance of the components can be made larger.

[0062] The herringbone grooves are made up of two types of spiralgrooves for pumping in and pumping out purposes. In other words,pumping-out spiral grooves 20 are positioned on the inner peripheralside, while pumping-in spiral grooves 21 are arranged on the outerperipheral side, with the crown 19 for making the bearing clearanceminimum positioned therebetween. The number of grooves per one round,the inclination angle of the grooves, and other features of the groovescan suitably be set according to their purposes.

[0063]FIG. 8(a) shows the pressure distribution observed during theoperation of the bearing having the above-described grooves. The y-axis73 indicates axial coordinates, while the x-axis 74 represents pressure.Reference numerals 75, 76, 77, 78, and 79 represent mean pressure valuesin a circumferential direction at respective axial positions. Thedrawing shows the pressure distribution without the influence of theatmospheric pressure, and therefore the pressure 75 at an outerperipheral point is zero. The pressure increases as denoted by thereference numeral 76 because of the grooves 21, and becomes constant inthe central region as indicated by the reference numeral 77. Thepressure decreases at a position where the grooves 20 are formed asindicated by the reference numeral 78. At the top 14 of the cone, thepressure is slightly higher than the atmospheric pressure as indicatedby the reference numeral 79.

[0064] The attitude of the rotary section is basically maintained by thehigh pressure 77 in the central region. A more specific account of theposition adjusting mechanism will be given below with reference to FIG.8(b). The pressure values 75, 76, 77, 78, and 79 in the pressuredistribution of FIG. 8(a) are mean values in circumferential directionsand they may locally vary if the sleeve 12 comes off-center or tiltswith respect to the shaft 11. FIG. 8(b) illustrates a state wherein thesleeve 12 is rotating as being inclined leftward at the upper partthereof and rightward at the lower part thereof with respect to theshaft 11. The load capacity, created by the grooves 20 in the centralregion where the clearance is made small by the crown 19, becomes unevenin the circumferential direction, i.e., the load capacity F11 on theright side becomes larger than the load capacity F12 on the left sidebecause the bearing clearance is smaller on the right side. Similarly,the pressure developed by the grooves 21 becomes uneven, the loadcapacity F21 on the right side being smaller than the load capacity F22on the left side where the bearing clearance is smaller. Here, the loadcapacity acts on the upper part of the sleeve 12 as moment ofL1*(F11-F12), while it acts on the lower part of the sleeve 12 as momentof L2*(F21-F22), where L1, L2 are the distances from an imaginaryfulcrum 66 corresponding to the cone apex and the respective points fromwhich the load capacity F11, F12, F21, F22 acts. The moment acts aroundthe fulcrum 66 as a force to make the bearing clearance at respectivepoints equal. It should be noted that the description given above issimplified and the moment actually counterbalances each other at allcircumferential and axial points, not only on the left and right sides.

[0065] In this way, by arranging a series of herringbone grooves on theconical surface with a small clearance region therebetween, a momentforce is generated that acts on the rotary section to equalize the upperand lower clearances between the shaft 11 and the sleeve 12, therebyadjusting the rotating attitude of the rotary section. Thus theprecession is further restricted in the fluid dynamic bearing motor ofthis embodiment. When the sleeve 12 comes off center with respect to theshaft 11, the pressure in the lubricant locally increases because of thewedge effect in the intermediate small-clearance band region formed bythe crown 19. A delay from the time when the bearing clearance isreduced until the time when a large pressure is developed may inducehalf whirls or other unstable movements of the rotary section. This iswhy the circumferential groove 40 is provided, as it distributes thelocally collected lubricant in circumferential directions, therebyenhancing the position adjusting effect by the grooves and preventinghalf whirls.

[0066]FIG. 9(a) and FIG. 9(b) illustrate the bearing section havingspiral grooves formed on the conical taper surface of the shaft 11. Theconical shaft 11 has a crown 19 so that the clearance between itsintermediate band region and the sleeve 12 becomes minimum. The spiralgrooves 22 for the pumping-in purpose are provided on the surface on theouter peripheral side of the shaft 11. Reference numerals 80, 81, and 82denote mean values of pressure at respective axial positions. As shown,the pressure becomes constant on the inner peripheral side from thespiral grooves 22 as indicated by the numeral 82. As can be seen fromFIG. 9(b), the pressure may vary in circumferential directions inaccordance with the change in the clearance between the shaft 11 and thesleeve 12 over the area from the grooves to the small-clearance bandregion. FIG. 9(b) illustrates the load capacity F21, F22 in a statewherein the sleeve 12 is tilted leftwards and the bearing clearance issmall on the lower left side. Since the load capacity is in inverseproportion to the bearing clearance, F22 is larger than F21. Thus, itacts on the sleeve 12 as moment of L2*(F21-F22), where L2 is thedistance from the imaginary fulcrum 66 conforming to the cone apex tothe point from which the load capacity F22 acts. The moment acts toequalize the bearing clearance, as a result of which the attitude of thesleeve 12 is adjusted. It should go without saying that the moment forceacts circumferentially on the sleeve 12, although the drawingillustrates moment forces acting from only both sides for the ease ofexplanation.

[0067] In this embodiment, even without the crown 19, whenever the shaftcomes off-center, the pressure distribution becomes uneven in thecircumferential direction, whereby the moment acts on the sleeve 12 toadjust its rotating attitude. However, the crown 19 causes the pressuredistribution to become uneven at a more peripherally outer position,whereby the moment force L2*(F21-F22) can be made larger.

[0068]FIG. 10(a) and FIG. 10(b) show the vicinity of the bearing sectionaccording to a further modified construction of the embodiment in whichgrooves are formed on both opposite surfaces of the bearing section.FIG. 10(b) is a cross-section of the shaft and the sleeve. The shaft 11has a permanent magnet 35 inside for generating magnetic attraction. Onits outer surface, a series of spiral grooves 20 is formed on its upperpart for the pumping-out purpose, and another series of spiral grooves21 is formed on its lower part for the pumping-in purpose. FIG. 10(a)shows a bearing surface of the sleeve 12 in a plan view. As shown, thesleeve 12 has on its bearing surface a plurality of herringbone grooves27 on its bearing surface. The grooves 20, 21, and 27 have a depth ofabout several micrometers, and grooves 20, 21 on the surface of theshaft 11 and those 27 on the sleeve 12 have different angular lengths inthe circumferential direction. In the specific example given in thesedrawings, the grooves 27 on the surface of the sleeve 12 have angularlengths of less than half as large as that of the grooves 20, 21 on theshaft 11 in the circumferential direction. The arrows 29, 30 indicatethe direction in which the sleeve 12 rotates.

[0069] Grooves pump the lubricant when the bearing rotates to increasethe pressure in the lubricant. The increased pressure, which issubstantially in inverse proportion to the bearing clearance, causes aforce to act on the rotary section to adjust its rotating attitude.Since the grooves are arranged at circumferentially spaced positions,even if the sleeve comes off-center with respect to the shaft and thebearing clearance becomes locally small, there is a delay until thebalance in the circumferential pressure distribution is disturbed. Thisdelay or time lag is in proportion to the angular length of the groovesin the circumferential direction. It is known that control systems withthe time lag between the change in the controlled variable and thecontrol over the change are susceptible to a resonant phenomenon, which,in the case of the fluid dynamic bearing, takes the form of precession,oil whip or other unstable movements.

[0070] In order to avoid such unstable movements, for example, thecircumferential length of the grooves 21 may be varied so that the timelag is varied. However, if the angular lengths of only several groovesin one round are changed, the possibility of the position adjustingforce not acting evenly increases, or other problem may arise.Therefore, in this embodiment, the grooves on the shaft 11 and those onthe sleeve 12 are varied in their angular lengths in the circumferentialdirection so as to both achieve the circumferential evenness in theposition adjusting force which is created by the increased pressure inlubricant, and the variety in the angular length of the grooves in thecircumferential direction. Machining of the grooves is generally noteasy and forming them on both bearing surfaces may lead to an increasein cost. However, the conical shaft 11 and the sleeve 12 in thisembodiment can both be produced by molding, and therefore such groovescan be provided without increasing cost. Thus a fluid dynamic bearingmotor with limited precession is realized.

[0071]FIG. 11 shows another modified construction of the embodimenthaving a channel 34 that runs through the shaft 11 from its truncatedcone top 14 to the outer periphery thereof. The channel 34 is providedfor circulating the lubricant compressed towards the top 14 of the shaft11 to the outside of the cone. The channel 34 is filled with fibrous orporous material to adjust the flow resistance such that pressure remainsat the top 14 of the cone, whereby the sleeve 12 can fly up swiftly atthe time of start-up, and whereby shock-absorbing effects are achievedbecause of the compressed lubricant that escapes and adjusts the dampinglevel. Moreover, galls produced on the sliding parts can be removed withthe structure of this example.

[0072]FIG. 12(a) and FIG. 12(b) illustrate a modified construction ofthe embodiment wherein the ring-shaped member can be adjusted in axialdirections. FIG. 12(b) is a cross sectional view of the bearing section,and FIG. 12(a) is an enlarged cross sectional view of part 89 of thering-shaped member and other components. In this example, the annularwall 23 has a protrusion 86 on its upper end, while the ring-shapedmember 24 has a corresponding through hole to match this protrusion. Thering-shaped member 24 is preliminarily coupled into the annular recess26 around the sleeve 12 and assembled to the shaft 11. Access holes 25are provided, through which the protrusion 86 and the through hole ofthe ring-shaped member 24 are engaged with each other. Then, using a jig88, the inner periphery of the ring-shaped member 24 is abutted onto theend face 87 of the annular recess 26. The ring-shaped member is thuscoupled to the protrusion 86 as being resiliently deformed.

[0073] In this assembling process, the ring-shaped member 24 isresiliently deformed in an axial direction by about 20 micrometers,while being coupled to the protrusion 86 firmly. Thereby, axialdisplacement of the rotary section including the hub 41 is restricted tobe about 20 micrometers even if it is subjected to large shocks. In thecase of hard disk drives, there is a strong demand for restricting axialdisplacement of the magnetic disk to a minimum By utilizing resilientdeformation of the ring-shaped member 24 as in this embodiment, suchrequirements can be met without higher demands for the tolerance ofvarious components. Alternatively, the ring-shaped member 24 and theprotrusion 86 may be joined after the assembly by bonding or welding tohave a higher strength to withstand large impacts.

[0074]FIG. 13 is a detailed cross sectional view of the top end of theshaft 11 and the sleeve 12 according to another modified construction ofthe embodiment, wherein conductive magnetic powder is mixed in thelubricant. The drawing is given in explanation of how the magneticpowder gathers in the magnetic field between the top end of the shaft 11and the sleeve 12 so as to achieve electrical conduction between therotary section and the fixed section. In the center at the inside toplimit of the sleeve 12 is provided a boss 36. Thus, the conductivemagnetic powder 37 mixed in the lubricant builds up between the centralboss 36 in the sleeve 12 and the magnet 35 where the magnetic flux isintense, thereby bridging the two and forming a conductive paththerebetween. The conductive magnetic powder 37 may be, for example,fine particles of ferrite magnet having a size of about 0.2 to 0.3micrometers so as to remain magnetic, with a coating of about 100angstrom thick gold.

[0075] The embodiments shown in FIG. 6, FIG. 11 employ a constructionwherein no weld joints are formed between the members in a portion whichthe lubricant contacts. In the prior art, separate components werejoined to provide a seal by swaging, bonding, or laser welding, but thiswas a major cause of later leakage of lubricant because of frequent bondfailure, leading to a fatal fault. The present invention provides afluid dynamic bearing motor free of the risk of oil leakage, as iteliminates joints in an area where the lubricant flows as shown in thisembodiment.

[0076] For the material of the bearing components such as shaft andsleeve, any of the metal materials such as stainless steel or copperalloy which have commonly been used for the fluid dynamic bearing can beused. Preferably, a thin film of nickel, titanium, diamond-like-carbon,or molybdenum disulfide should be formed on one of the conical tapersurfaces, so as to decrease the friction at the time of starting up andstopping the motor.

[0077] Regarding the manufacturing method of the bearing components, notto mention the shaft having a convex shape, the sleeve, although havinga concave shape, it can be easily released, because its tapered top isopened. Therefore they both can be formed at one time including thegrooves, by any known techniques such as press molding or injectionmolding. Accordingly, the bearing components can also be made of aceramics or sintered alloy by molding, or of a resin material havingsuperior antifriction properties such as polyphenyl sulfide resin (PPS)containing carbon fiber by molding, whereby a reduction in productioncost is achieved.

[0078] Although the embodiments shown in above have been described ashaving the sleeve 12 or the shaft 11 and the hub 41 formed in one piece,they may be separate components and assembled together. Whether theyshould be produced in one piece or separately may be determined case bycase so that the cost is lower, taking into consideration thecharacteristics and specifications required for each component. In theapplication of the invention to a hard disk drive as has been shown inthese embodiments, however, there are stringent specifications withregard to the height and tilt of the install surface of the magneticdisk. Since these are strongly affected by their positional relationshipwith the bearing surface, it is more preferable to form the sleeve 12 orthe shaft 11 and the hub 41 in one piece to achieve higher precision.The fluid dynamic bearing motor according to the present inventionenables the integral structure of the sleeve or the shaft and the huband realizes a high-precision, low-cost motor.

[0079] According to a fluid dynamic bearing motor of the presentinvention, the bearing section has a simple structure wherein groovesare formed on a conical taper surface for increasing the pressure inlubricant and creating a load capacity, which is balanced with magneticattraction. With this structure, the attitude of the rotary section inthe bearing is made stable, and a reliable seal of the lubricant isachieved even in high-speed operation. The bearings can be mass-producedat low cost by molding, and the total thickness of the motor can bereduced. Further, a temperature compensation of the load capacity forsupporting the rotary section is achieved, and the current required foroperating the motor is reduced. Therefore, the motor according to theinvention is particularly suitable for small, rotary memory devices suchas magnetic or optical disk devices, or cooling fans for CPUs.

[0080] While there has been described what are at present considered tobe preferred embodiments of the present invention, it will be understoodthat various modifications may be made thereto, and it is intended thatthe appended claims cover all such modifications as fall within the truespirit and scope of the invention.

What is claimed is:
 1. A fluid dynamic bearing motor, comprising: ashaft having a diminishing conical taper surface; a sleeve having aconical concavity opposite the shaft; lubricant filled in a clearancebetween the shaft and the sleeve; and means for generating magneticattraction between one end of the shaft and a cone apex of the sleeve,including a magnet and a magnetic material, the magnet being provided inone of the shaft and the sleeve and the magnetic being provided in theother one thereof; and an annular wall arranged around the shaft to facean outer circumferential wall of the sleeve, a clearance between theannular wall and the outer circumferential wall of the sleeve beingincreased in width toward an open end to form a taper seal of thelubricant; wherein a plurality of grooves are formed on at least aconical taper surface of one of the shaft and the sleeve, the groovesbeing provided for creating load capacity when the motor rotates,whereby rotating parts of the motor are supported by axial components ofthe load capacity balanced with said magnetic attraction.
 2. The fluiddynamic bearing motor according to claim 1, wherein the magneticattraction generating means is composed of a permanent magnet heldwithin the shaft and a magnetic material provided in the sleeve, andwherein said permanent magnet is initially held movably but firmlyenough to overcome said magnetic attraction, and is adjusted and fixedin position at the time of assembling such that the permanent magnetmakes contact with the sleeve when the motor is stationary, while theyare brought out of contact when the motor is rotating, by a distanceequal to or shorter than an axial flying height determined on conicalsurfaces of the shaft and sleeve.
 3. The fluid dynamic bearing motoraccording to claim 1, wherein the magnetic attraction generating meansis composed of a permanent magnet held within the shaft and a magneticmaterial provided in the sleeve, the permanent magnet being assembledwith the shaft such that it is initially held movably but firmly enoughto overcome said magnetic attraction as being substantially protrudedfrom one end of the shaft, and is pressed into the shaft by a pressurelarger than the magnetic attraction applied from both ends of the shaftand the sleeve to a predetermined position, where the cone apex of thesleeve or a plate spring interposed between the apex of the sleeve andthe permanent magnet is resiliently deformed, whereby when the motor isstationary the permanent magnet and the apex of the sleeve or the platespring make contact with each other, while they are brought out ofcontact when the motor is rotating, by a distance equal to or shorterthan an axial flying height determined on conical surfaces of the shaftand sleeve.
 4. The fluid dynamic bearing motor according to claim 1,further including a crown of about several micrometers provided on theconical taper surface of one of the shaft and the sleeve so as to makethe clearance between the opposite taper surfaces of the shaft and thesleeve be minimum at an axially intermediate region, wherein the groovesare spiral grooves and provided on one or both sides of said axiallyintermediate region where the clearance between the shaft and the sleeveis minimum, for pumping the lubricant towards said intermediate region.5. The fluid dynamic bearing motor according to claim 4, furtherincluding a circumferential groove provided on the conical taper surfaceof at least one of the shaft and the sleeve where the clearancetherebetween is minimum because of the crown.
 6. The fluid dynamicbearing motor according to claim 1, wherein the grooves are formed onboth opposite conical taper surfaces of the shaft and the sleeve ataxially opposite positions, the grooves having different angular lengthsfrom each other in a circumferential direction.
 7. The fluid dynamicbearing motor according to claim 1, further including a ring-shapedmember fixed to one end of said annular wall, and an annular recessprovided in the outer circumferential wall of the sleeve and makingengagement with the ring-shaped member so as to restrict axial movabledistance of the rotating parts.
 8. The fluid dynamic bearing motoraccording to claim 7, wherein the ring-shaped member is fixed to one endof the annular wall by any one of interfitting, bonding, and welding,access holes being provided in either one of a fixed member and a rotarymember opposite said one end of the annular wall for enabling the fixingof the ring-shaped member to be performed.
 9. The fluid dynamic bearingmotor according to claim 8, further including a means for establishingcoupling engagement between the ring-shaped member and one end of theannular wall, wherein the ring-shaped member is coupled to one end ofthe annular wall with an inner peripheral portion thereof being pressed,through the access hole, to be resiliently deformed and abutted onto anend face of said annular recess, whereby said resilient deformation ofthe ring-shaped member determines a permissible range of the axialdisplacement of the rotating parts.
 10. The fluid dynamic bearing motoraccording to claim 1, wherein the lubricant is a magnetic oil, wherebyair bubbles are readily removed from the lubricant.
 11. The fluiddynamic bearing motor according to claim 1, wherein conductive magneticpowder is mixed in the lubricant for bridging one end of the shaft andthe apex of the sleeve to form an electrically conductive paththerebetween.